1. Field of the Invention
The present invention relates to an improvement in a continuously variable transmission apparatus incorporating a toroidal-type continuously variable transmission unit which is used as a vehicle (automotive) automatic transmission apparatus, and more particularly to an improvement in characteristic when a vehicle is stationary or drives at extremely low speeds.
2. Description of Related Art
The usage of a toroidal-type continuously variable transmission unit as shown in FIGS. 4 to 6 as an automotive automatic transmission apparatus has been studied and has now been implemented partially. This toroidal-type continuously variable transmission unit is such as to be called a double-cavity type, in which input side disks 2, 2 are supported around both end portions of an input shaft 1 via ball splines 3, 3. Consequently, these two input side disks 2, 2 are supported coaxially with each other and rotatably in a synchronous fashion. In addition, an output gear 4 is supported around an intermediate portion of the input shaft 1 in such a manner as to freely rotate relative to the input shaft 1. Then, output side disks 5, 5 are brought into spline engagement with both end portions of a cylindrical portion provided in a central portion of the output gear 4, respectively. Consequently, both the output side disks 5, 5 rotate together with the output gear 4 in a synchronous fashion.
In addition, a plurality (normally, two to three) of power rollers 6, 6 are interposed between the input side disk 2, 2 and the output side disk 5, 5, respectively. The power rollers 6, 6 are supported rotatably on inner surfaces of trunnions 7, 7 via support shafts 8, 8 and a plurality of rolling bearings, respectively. The trunnions 7, 7 each freely swing to be displaced about pivot shafts 9, 9 which are provided on longitudinal (in a vertical direction in FIGS. 4 and 6, and in a direction normal to the surface of the piece of paper on which FIG. 5 is illustrated) end portions of the trunnions 7, 7 coaxially therewith, respectively. While operations of inclining the respective trunnions 7, 7 are implemented by displacing the trunnions 7, 7 in axial directions of the pivot shafts 9, 9 by hydraulic actuators 10, 10, inclination angles of all the trunnions 7, 7 are synchronized with each other hydraulically and mechanically.
Namely, when changing the inclination angles of the respective trunnions 7, 7 in order to change the transmission ratio between the input shaft 1 and the output gear 4, the respective trunnions 7, 7 are displaced in opposite directions by the respective actuators 10, 10, for example, the right-hand side power roller 6 in FIG. 6 being displaced downwardly in the same drawing and the left-hand side power roller 6 in FIG. 6 being displaced upwardly in the same drawing, respectively. As a result, the orientations of tangential forces change which act on abutment portions between circumferential surfaces of the respective power rollers 6, 6 and inner surfaces of the respective input side disks 2, 2 and the respective output side disks 5, 5 (a side slip is generated at the abutment portions). Then, in association with the change in orientation of the forces, the respective trunnions 7, 7 swing (incline) in the opposite directions to each other about the pivot shafts 9, 9 pivotally supported on support plates 11, 11. As a result, the abutment positions between the circumferential surfaces of the respective power rollers 6, 6 and the inner surfaces of the input and output side disks 2, 5 change, and the rotational transmission ratio between the input shaft 1 and the output gear 4 changes.
The supply and discharge of hydraulic fluid to and from the respective actuators 10, 10 are implemented by a single control valve 12 irrespective of the number of actuators 10, 10, and it is designed that the movement of any one of the trunnions 7, 7 is fed back to the control valve 12. The control valve 12 has a sleeve 14 adapted to be displaced axially (in left and right directions in FIG. 6, and in directions normal to the surface of the piece of paper on which FIG. 4 is illustrated) by a stepping motor 13 and a spool 15 which is fittingly installed on an inside-diameter side of the sleeve 14 in such a manner as to be displaced axially. In addition, a precess cam 18 is fixed to an end portion of either of rods 17, 17 attached to the trunnions 7, 7 to connect the respective trunnions 7, 7 with pistons 16, 16 of the respective actuators 10, 10, whereby a feedback mechanism is configured in which the movement of the rod 17, that is, a resultant value of an axial displacement amount and a rotational displacement amount thereof is transmitted to the spool 15 via the precess cam 18 and a link arm 19. In addition, a synchronous cable 20 is provided so as to extend between the respective trunnions 7, 7 so that the inclination angles of the respective trunnions 7, 7 can also mechanically be synchronized with each other should the hydraulic system fail.
In changing the transmission ratio, the sleeve 14 is displaced by the stepping motor 13 to a predetermined position which matches a transmission ratio attempted to be obtained to thereby open a flow path of the control valve 12 in a predetermined direction. As a result, the hydraulic fluid is sent into the respective actuators 10, 10 in a predetermined direction, whereby the actuators 10, 10 displaces the trunnions 7, 7 in a predetermined direction, respectively. Namely, as the hydraulic fluid is sent in, the respective trunnions 7, 7 are displaced in the axial directions of the respective pivot shafts 9, 9 and swing about the pivot shafts 9, 9, respectively. Then, the movement (the axial and swinging displacements) of any one 7 of the trunnions is transmitted to the spool 15 via the precess cam 18 fixed to the end portion of the rod 17 and the link arm 19 to thereby displace the spool 15 axially. As a result, with the trunnion 7 being displaced a predetermined amount, the flow path of the control valve 12 is closed, and the supply and discharge of hydraulic fluid to and from the respective actuators 10, 10 are stopped.
The movement of the control valve 12 based on the displacements of the trunnion 7 and a cam surface 21 of the precess cam 18 as the aforesaid occurs will be as follows. Firstly, when the trunnion 7 is displaced axially as the flow path of the control valve 12 is opened, as has been described previously, the trunnion 7 starts to swing to be displaced about the pivot shafts 9, 9 by virtue of the side slip generated at the abutment portions between the circumferential surface of the power roller 6 and the inner surfaces of the input side disk 2 and the output side disk 5. In addition, the displace of the cam surface 21 in association with the axial displacement of the trunnion 7 is transmitted to the spool 15 via the link arm 19, and the spool 15 is then displaced axially to thereby change the changeover state of the control valve 12. To be specific, the control valve 12 changes over in a direction in which the trunnion 7 is returned to a neutral position by the actuator 10.
Consequently, the trunnion 7 starts to be displaced in the opposite direction so as to be returned to the neutral position immediately the trunnion 7 is displaced axially. However, as long as the trunnion 7 needs to be displaced from the neutral position, the trunnion 7 continues to swing about the respective pivot shafts 9, 9. As a result, a displacement with respect to a circumferential direction of the cam surface 21 of the precess cam 18 is transmitted to the spool 15 via the link arm 19, whereby the spool 15 is displaced axially. Then, at the same time that the trunnion 7 returns to the neutral position with the inclination angle of the trunnion 7 having reached the predetermined angle that matches the transmission ratio attempted to be obtained, the control valve 12 is closed, and the supply and discharge of hydraulic fluid to and from the actuator 10 is stopped. As a result of this, the inclination angle of the trunnion 7 becomes an angle which matches an amount in which the sleeve 14 is displaced axially by the stepping motor 13.
When the toroidal-type continuously variable transmission unit described above is operated, one of the input side disks 2 (the left-hand side disk in FIGS. 4, 5) is rotatably driven by a drive shaft 22 connecting to a drive source such as an engine via a loading cam-type pressing device 23 which is shown in the figures. As a result, the pair of input side disks 2, 2 which are supported on the end portions of the input shaft 1 rotate synchronously while being pressed against in directions in which they approach each other. Then, the rotations are transmitted via the power rollers 6, 6 to the output side disks 5, 5, respectively, so as to be taken from the output gear 4.
When the power is transmitted from the input side disks 2, 2 to the output side disks 5, 5, respectively, as has been described above, as friction is generated at rolling contact portions (traction portions) between the circumferential surfaces of the power rollers 6, 6 supported on the inner surfaces of the trunnions 7, 7 and the inner surfaces of the respective disks 2, 5, axial forces of the pivot shafts 9, 9 provided on the end portions of the respective trunnions 7, 7 are applied to the trunnions 7, 7, respectively. This force is referred to as a so-called 2 Ft, and the magnitude thereof is proportional to a torque transmitted from the respective input side disks 2, 2 to the respective output side disks 5, 5 (or, from the respective output side disks 5, 5 to the respective input side disks). Then, the force 2 Ft is borne by the respective actuators 10, 10. Consequently, in an operation of the toroidal-type continuously variable transmission unit, a difference in pressure between pairs of hydraulic pressure chambers 24a, 24b existing on both sides of the pistons 16, 16 which constitute the respective actuators 10, 10 is proportional to the magnitude of the force 2 Ft.
When changing the rotational speeds between the input shaft 1 and the output gear 4, firstly in the event that a reduction in speed is implemented between the input shaft 1 and the output gear 4, the respective trunnions 7, 7 are moved in the axial directions of the respective pivot shafts 9, 9 by the respective actuators 10, 10, and the respective trunnions 7, 7 are caused to swing to positions shown in FIG. 5. Then, as shown in FIG. 5, the circumferential surfaces of the respective power rollers 6, 6 are brought into abutment with the inner surfaces of the respective input side disks 2, 2 at portions closer to the centers thereof and the inner surfaces of the respective output side disks 5, 5 at portions closer to the outer circumferences thereof. On the contrary, in the event that an increase in speed is implemented, the respective trunnions 7, 7 are caused to swing in opposite directions to those shown in FIG. 5, and the respective trunnions 7, 7 are inclined so that the circumferential surfaces of the respective power rollers 6, 6 are brought into abutment with the inner surfaces of the input side disks 2, 2 at portions closer to the outer circumferences thereof and the inner surfaces of the output side disks 5, 5 at portions closer to the centers thereof. In case the inclination angles of the respective trunnions 7, 7 are made to be intermediate, an intermediate transmission ratio (speed ratio) can be obtained between the input shaft 1 and the output gear 4.
Furthermore, when incorporating the toroidal-type continuously variable transmission unit which is constructed and which operates as has been described above in an actual automotive continuously variable transmission unit, it has conventionally been proposed that the toroidal-type continuously variable transmission unit is combined with a gear-type differential unit such as using a planetary gear mechanism so as to constitute a continuously variable transmission apparatus. For example, described in U.S. Pat. No. 6,251,039 is a continuously variable transmission apparatus in which the rotational state of an output shaft is changed over from forward to backward rotations or vice versa with a stationary state being interposed therebetween with an input shaft being kept rotating in one direction, this being referred to as a so-called geared neutral. FIG. 7 illustrates the continuously variable transmission apparatus described in the U.S. Pat. No. 6,251,039. This continuously variable transmission apparatus is constituted by a toroidal-type continuously variable transmission unit 25 and a planet gear-type transmission unit 26. Of the two transmission units, the toroidal-type continuously variable transmission unit 25 includes an input shaft 1, a pair of input side disks 2, 2, an output side disk 5a and a plurality of power rollers 6, 6. In the illustrated example, the output side disk 5a is such as to be constructed by allowing outer surfaces of a pair of output side disks to abut each other so that the pair of output side disks become integrated as a whole.
In addition, the planet gear-type transmission 26 includes a carrier 27 which is fixedly connected to the input shaft 1 and one (the right-hand side one in FIG. 7) of the input side disks 2. A primary transmission shaft 29 having planet gear elements 28a, 28b fixedly provided at end portions thereof is rotatably supported on the carrier 27 at a radially intermediate portion thereof. In addition, a secondary transmission shaft 31 having sun gears 30a, 30b fixedly provided at end portions thereof is supported rotatably and coaxially with the input shaft 1 on an opposite side to the input shaft 1 with the carrier 27 being held therebetween. Then, the respective planet gear elements 28a, 28b are brought into mesh engagement, respectively, with a sun gear 33 fixedly provided on a distal end portion (a right end portion in FIG. 7) of a hollow rotational shaft 32 which is connected to the output side disk 5a at a proximal end portion (a left end portion in FIG. 7) thereof and the sun gear 30a fixedly provided on one end portion (left end portion in FIG. 7) of the secondary transmission shaft 31. In addition, one of the planet gear elements (the left-hand side element) 28a is brought into mesh engagement with a ring gear 35 which is provided in such a manner as to freely rotate around the carrier 27 via another planet gear element 34.
On the other hand, planet gear elements 37a, 37b are rotatably supported on a second carrier 36 provided around the sun gear 30b fixedly provided at the other end portion (the right end portion in FIG. 7) of the secondary transmission shaft 31. Note that the second carrier 36 is fixedly provided at a proximal end portion (a left end portion in FIG. 7) of an output shaft 38 which is disposed coaxially with the input shaft 1 and the secondary transmission shaft 31. In addition, the respective planet gear elements 37a, 37b mesh with each other, and one of the planet gear elements 37a meshes with the sun gear 30b, whereas the other planet gear element 37b meshes with a second ring gear 39 which is provided around the second carrier 36 in such a manner as to freely rotate therearound. Additionally, the ring gear 35 and the secondary carrier 36 are made to engage with and disengage from each other by a low-speed clutch 40, and the secondary ring gear 39 and a fixed portion such as a housing are made to engage with and disengage from each other by a high-speed clutch 41.
In the case of the continuously variable transmission apparatus shown in FIG. 7 which has been described heretofore, in a so-called low speed mode in which the low-speed clutch is engaged, whereas the high-speed clutch 41 is disengaged, the power of the input shaft 1 is transmitted to the output shaft 38 via the ring gear 35. Then, the transmission ratio of the whole continuously variable transmission unit, that is, the transmission ratio between the input shaft 1 and the output shaft 38 is changed by changing the transmission ratio of the toroidal-type continuously variable transmission unit 25. In the low-speed mode like this, the transmission ratio of the whole continuously variable transmission apparatus changes infinitely. Namely, by regulating the transmission ratio of the toroidal-type continuously variable transmission unit 25, with the input shaft 1 being kept rotating in one direction, the rotational state of the output shaft 38 can freely be converted from forward to backward rotations or vice versa with the stationary state being held therebetween.
During an acceleration in the low-speed mode like this or constant speed running, a torque (a passing torque) passing through the toroidal-type continuously variable transmission unit 25 is applied from the input shaft 1 to the output side disk 5a via the carrier 27 and the primary transmission shaft 29, the sun gear 33, and the hollow rotational shaft 32, and is then applied from this output side disk 5a to the respective input side disks 2, 2 via the respective power rollers 6, 6. Namely, the torque passing through the toroidal-type continuously variable transmission unit 25 during the acceleration or constant speed running of the engine circulates in a direction in which the respective input side disks 2, 2 receive the torque from the respective power rollers 6, 6.
In contrast to this, in a so-called high-speed mode in which the low-speed clutch 40 is disengaged, whereas the high-speed clutch 41 is engaged, the power of the input shaft 1 is transmitted to the output shaft 38 via the primary and secondary transmission shafts 29, 31. Then, the transmission ratio of the whole continuously variable transmission apparatus is changed by changing the transmission ratio of the toroidal-type continuously variable transmission unit 25. In this case, the greater the transmission ratio of the toroidal-type continuously variable transmission unit 25 becomes, the greater the transmission ratio of the whole continuously variable transmission apparatus becomes.
Note that a torque passing through the toroidal-type continuously variable transmission unit 25 during an acceleration in the high-speed mode like this or constant speed running is applied in a direction in which the respective input side disks 2, 2 apply the torque to the respective power rollers 6, 6.
For example, in the case of a continuously variable transmission apparatus which has the construction illustrated in FIG. 7 and which can realize a so-called infinite transmission ratio in which the output shaft 38 is stopped while the input shaft 1 is allowed to rotate, with a view to securing the durability of and ensuring the facilitation of operation of the toroidal-type continuously variable transmission unit 25, it is important to maintain a torque applied to the toroidal-type continuously variable transmission unit 25 at an appropriate value in a state in which the transmission ratio is extremely increased with a state being involved in which the output shaft 38 is stopped. This is because, as is clear from a relationship of “rotational driving force=rotational speed×torque”, in a state in which the transmission ratio is extremely large and the output shaft 38 is stopped or rotates at extremely low speeds with the input shaft 1 being kept rotating, the torque (the passing torque) passing through the toroidal-type continuously variable transmission unit 25 becomes larger than a torque applied to the input shaft 1. Due to this, in order to secure the durability of the toroidal-type continuously variable transmission unit 25 without increasing the size thereof, there is produced a necessity of implementing a strict control to maintain the torque at the appropriate value as has been described above. To be specific, in order to stop the output shaft 38, while making the torque inputted into the input shaft 1 as small as possible, a control including the drive source is required.
In addition, in the state in which the transmission ratio is extremely large, even if the transmission ratio of the toroidal-type continuously variable transmission unit 25 changes slightly, the torque applied to the output shaft 38 changes largely. Due to this, unless the regulation of the transmission ratio of the toroidal-type continuously variable transmission unit 25 is implemented strictly, the driver possibly feels an abnormal feeling or the proper operation of the vehicle is possibly made difficult. For example, in the case of an automotive automatic transmission apparatus, a stationary state of the vehicle is maintained with a brake pedal being depressed by the driver. In this case, unless a strict regulation of the transmission ratio of the toroidal-type continuously variable transmission unit 25 is implemented and, as a result, a large torque is applied to the output shaft 38, a large effort to depress the brake pedal is required to stop the vehicle, whereby the driver is fatigued largely. On the contrary, unless a strict regulation of the transmission ratio of the toroidal-type continuously variable transmission unit 25 is implemented at the start of the vehicle, and, as a result, the torque applied to the output shaft 38 becomes too small, a smooth start of the vehicle is not possibly attained, or the vehicle is possibly reversed when attempted to be started upwardly from the rest on a slope. Consequently, when the vehicle is being stopped or is driving at extremely low speeds, not only does the torque that is transmitted from the drive source to the input shaft 1 need to be controlled, but also the regulation of the transmission ratio of the toroidal-type continuously variable transmission unit 25 needs to be implemented strictly.
In view of the points raised above, JP-A-10-103461 describes a construction which regulates a torque (a passing torque) passing through a toroidal-type continuously variable transmission unit by directly controlling a difference in pressure between hydraulic actuator portions for displacing trunnions.
In the case of the construction such as described in JP-A-10-3461, however, because the control is implemented based only on the difference in pressure, it is difficult to make the posture of the trunnion stationary at the moment the passing torque has matched a target value. Specifically speaking, since an amount in which the trunnion is displaced becomes large due to the control of the torque, a so-called overshoot (and, furthermore, a hunting in association with such an overshoot) tends to occur easily in which the trunnion does not stop at the moment the passing torque has matched the target value but continues to be displaced, and hence the control of the passing torque is not stabilized.
In particular, in the case of the toroidal-type continuously variable transmission unit 25 having no so-called cast angle such as the general half toroidal-type continuously variable transmission illustrated in FIGS. 4 to 6 in which the directions of the pivot shafts 9, 9 provided at the end portions of the trunnions 7, 7 and the direction of the central axis of the input and output side disks 2, 5 become normal to each other, the aforesaid overshoot is easy to occur. In contrast to this, in the case of a construction having the cast angle such as a general full toroidal-type continuously variable transmission unit, since a force is applied in a direction in which the overshoot is converged, even with the construction such as illustrated in JP-A-103461, it is considered that a sufficient torque control can be implemented.
In view of these situations, even with a continuously variable transmission apparatus which incorporates therein a toroidal-type continuously variable transmission unit having no cast angle such as the general half toroidal-type continuously variable transmission unit, an example of the construction of a continuously variable transmission apparatus is shown in FIG. 8 which can implement a strict control over a torque which passes through the toroidal-type continuously variable transmission. While the continuously variable transmission apparatus shown in FIG. 8 has a similar function to that of the conventionally known continuously variable transmission apparatus shown in FIG. 7, by devising the construction of the part thereof where the planet gear-type transmission 26a is provided, the assembling property of the part thereof where the planet gear-type transmission 26a is provided is improved.
Primary and secondary planetary gears 42, 43, which are of a double-pinion type, respectively, are supported, respectively, on sides of a carrier 27a which rotates together with an input shaft 1 and a pair of input side disks 2, 2. Namely, these primary and secondary planetary gears 42, 43 include a pair of planet gear elements 44a, 44b and a pair of planet gear elements 45a, 45b, respectively. Then, while the planet gear elements 44a and 44b, and 45a and 45b are made to mesh with each other, respectively, the radially inward planet gear elements 44a, 45a are made to mesh, respectively, with primary and secondary sun gears 47, 48 which are fixedly provided at a distal end portion (a right end portion in FIG. 8) of a hollow rotational shaft 32a which is connected to an output side disk 5a at a proximal end portion (a left end portion in FIG. 8) thereof and at one end portion (a left end portion in FIG. 8) of a transmission shaft 46, respectively, and the radially outward planet gear elements 44b, 45b are made to mesh with a ring gear 49, respectively.
On the other hand, planet gear elements 51a, 51b are rotatably supported on a secondary carrier 36a provided around a tertiary sun gear 50 fixedly provided at the other end portion (a right end portion in FIG. 8) of the transmission shaft 46. Note that this secondary carrier 36a is fixedly provided at a proximal end portion (a left end portion in FIG. 8) of an output shaft 38a disposed concentrically with the input shaft 1. In addition, while the respective planet gear elements 51a, 51b are made to mesh with each other, the radially inward planet gear element 51a is made to mesh with the tertiary sun gear 50, whereas the radially outward planet gear element 51b is made to mesh with a secondary ring gear 39a provided rotatably around the second carrier 36a. Additionally, the ring gear 49 and the secondary carrier 36a are freely engaged with and disengaged from each other via a low-speed clutch 40a, whereas the secondary ring gear 39a and a fixed portion such as a housing are engaged with and disengaged from each other via a high-speed clutch 41a. 
In the case of the improved continuously variable transmission apparatus that is constructed as has been described above, in a state in which the low-speed clutch 40a is engaged, whereas the high-speed clutch 41a is disengaged, the power of the input shaft 1 is transmitted to the output shaft 38a via the ring gear 49. Then, by changing the transmission ratio of the toroidal-type continuously variable transmission unit 25, the speed ratio eCVT of the whole continuously variable transmission apparatus, that is, the speed ratio between the input shaft 1 and the output shaft 38a is changed. As this occurs, a relationship between the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 and the speed ratio eCVT of the whole continuously variable transmission apparatus will be expressed by the following expression (1), assuming that a ratio between the number of teeth m49 of the ring gear 49 and the number of teeth m47 of the primary sun gear 47 is i1 (=m49/m47),eCVT=(eCVU+i1−1)/i1  (1)
Then, in the event that the ratio i1 between the numbers of teeth of the ring gear 49 and the primary sun gear 47 is 2, a relationship between both the speed ratios eCVU, eCVT changes as illustrated by a line segment α in FIG. 9.
In contrast to this, in a state in which the low-speed clutch 40a is disengaged, whereas the high-speed clutch 41a is engaged, the power of the input shaft 1 is transmitted to the output shaft 38a via the primary planetary gear 42, the ring gear 49, the secondary planetary gear 43, the transmission shaft 46, the planet gear elements 51a, 51b, and the secondary carrier 36a. Then, by changing the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25, the speed ratio eCVT of the whole continuously variable transmission apparatus is changed. As this occurs, a relationship between the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 and the speed ratio eCVT of the whole continuously variable transmission apparatus will be expressed by the following expression (2). Note that, in this expression (2), i1 denotes a ratio (m49/m47) between the number of teeth m49 of the ring gear and the number of teeth m47 of the primary sun gear 47, i2 denotes a ratio (m49/m48) between the number of teeth m49 of the ring gear 49 and the number of teeth m48 of the secondary sun gear 48, and i3 denotes a ratio (m39/m50) between the number of teeth m39 of the secondary ring gear 39a and the number of teeth m50 of the tertiary sun gear 50, respectively.eCVT={1/(1−i3)}·{1+(i2/i1) (eCVU−1)}  (2)
Then, in the event that, of these ratios, i1 is 2, i2 is 2.2, and i3 is 2.8, the relationship between the speed ratios eCVU, eCVT changes as illustrated by a segment line β in FIG. 9.
In the case of the continuously variable transmission apparatus which is constructed and which functions as has been described above, as is clear from the segment line α in FIG. 9, a so-called infinite transmission ratio state is created in which the output shaft 38a is stopped while allowing the input shaft 1 to rotate. However, in the state like this in which the output shaft 38a is stopped or is allowed to rotate at extremely low speeds with the input shaft 1 being kept rotating, as has been described previously, the torque (the passing torque) passing through the toroidal-type continuously variable transmission unit 25 becomes larger than a torque that is applied from the engine which is the drive source to the input shaft 1. Due to this, when the vehicle is stationary or drives at extremely low speeds, the torque that is inputted from the drive source to the input shaft 1 needs to be regulated strictly so that the passing torque becomes too large (or too small).
In addition, during the driving at extremely low speeds, in a state close to the state in which the output shaft 38a is stopped, that is, a state in which the transmission ratio of the continuously variable transmission is very large and the rotational speed of the input shaft 38a is largely slow when compared with the rotational speed of the input shaft 1, a slight variation in transmission ratio of the continuously variable transmission apparatus largely varies a torque that is applied to the output shaft 38a. Due to this, in order to ensure a smooth driving operation, the torque that is inputted from the drive source to the input shaft 1 also needs to be regulated properly.
In addition, during an acceleration in the low-speed mode like this or constant speed running, as with the conventional structure shown in FIG. 7, the passing torque is applied from the input shaft 1 to the output side disk 5a via the carrier 27a and the primary planetary gear 42, the primary sun gear 47, and the hollow rotational shaft 32a, and is then applied from this output side disk 5a to the respective input side disks 2, 2 via the respective power rollers 6, 6 (refer to FIG. 7). Namely, during the acceleration or constant speed running, the passing torque circulates in a direction in which the respective input side disks 2, 2 receive the torque from the respective power rollers 6, 6.
Due to this, according to a method and apparatus for controlling the transmission ratio by the aforesaid structure, the torque that is inputted from the drive source to the input shaft 1 is designed to be regulated properly as illustrated in FIG. 10. Firstly, the rotational speed of the engine which is the drive source is controlled roughly. Namely, the rotational speed of the engine is regulated to a point a within a range w in FIG. 10. In conjunction with this, a transmission ratio of the toroidal-type continuously variable transmission unit 25 is set which is understood to be required so that the rotational speed of the input shaft 1 of the continuously variable transmission apparatus coincides with the controlled rotational speed of the engine. This setting job is implemented based on the aforesaid expression (1). Namely, it is when the so-called low-speed mode in which the low-speed clutch 40a is engaged, whereas the high-speed clutch 41a is disengaged is applied that the torque that is transmitted from the engine to the input shaft 1 is strictly controlled by the method above. Consequently, the transmission ratio of the toroidal-type continuously variable transmission unit 25 is set by the expression (1) above so that the rotational speed of the input shaft 1 become a value which corresponds to the required rotational speed of the output shaft 38a. 
In addition, differences in pressure between pairs of hydraulic pressure chambers 24a, 24b (refer to FIGS. 6, 12) which constitute hydraulic actuators 10, 10 for displacing trunnions 7, 7 disposed in the toroidal-type continuously variable transmission unit 25 in axial directions of pivot shafts 9, 9 are measured by a hydraulic pressure sensor 52 (refer to FIG. 2 that will be described later on). This hydraulic pressure measuring job is performed in a state in which the rotational speed of the engine is roughly controlled (however, the rotational speed should be maintained constant) and in accordance with this, the transmission ratio of the toroidal-type continuously variable transmission unit 25 is set based on the expression (1) as has been described above. Then, a torque (a passing torque) that passes through the toroidal-type continuously variable transmission unit 25 is calculated from differences in pressure so obtained based on the measuring job.
Namely, since the differences in pressure are proportional to a torque TCVU that passes through the toroidal-type continuously variable transmission unit 25 as long as the transmission ratio of the toroidal-type continuously variable transmission unit 25 remains constant, this torque TCVU can be obtained from the differences in pressure. As has been described previously, this is because the respective actuators 10, 10 bear the force of 2 Ft that has a magnitude proportional to a torque (=a torque TCVU that passes through the toroidal-type continuously variable transmission unit 25) that is transmitted from the input side disks 2, 2 to the output side disk 5a (or from the output side disk 5a to the input side disks 2, 2).
On the other hand, the torque TCVU is also obtained by the following expression (3).TCVU=eCVU·TIN/(eCVU+(i1−1) ηCVU)  (3)In this expression (3), eCVU denotes the transmission ratio of the toroidal-type continuously variable transmission unit 25, TIN denotes a torque that is inputted from the engine to the input shaft 1, i1 denotes a ratio of numbers of teeth (a ratio between the number of teeth m49 of the ring gear 49 and the number of teeth m47 of the primary sun gear 47) of the planetary gear transmission unit for the primary planetary gear 42, and ηCVU denotes the efficiency of the toroidal-type continuously variable transmission unit 25, respectively.
Then, a deviation ΔT (=TCVU1−TCVU2) between a torque TCVU1 actually passing through the toroidal-type continuously variable transmission unit 25 which is obtained from the difference in pressure and a target passing torque TCVU2 which is obtained from the expression (3) is obtained based on the actual passing torque TCVU1 and the target value TCVU2. Then, the speed ratio of the toroidal-type continuously variable transmission unit 25 is regulated in a direction that cancels the deviation ΔT (ΔT=0). Note that since the deviation ΔT between the torques and a deviation between the differences in pressure are in a proportional relation, the transmission ratio regulating job can be performed by either of the deviation between the pressures and the deviation between the differences in pressure. Namely, from the technical perspective, the transmission ratio control through the torque deviation is identical with the transmission ratio control through the differences-in-pressure deviation.
For example, a case is considered in which the torque TIN with which the engine drives the input shaft 1 drastically changes to rapidly decrease as the rotational speed of the input shaft 1 increases in the area where the torque TCVU1 (the measured value) that actually passes through the toroidal-type continuously variable transmission unit 25 is regulated to the target value TCVU2 as shown in FIG. 10. The property of the engine like this can easily be obtained even in a low rotational speed area provided that the engine is electronically controlled. In the event that, with the engine having such an engine property, when compared with the target value TCVU2, the measured value TCVU1 of the torque has similarly a deviation in a direction in which the respective input side disks 2, 2 receive the torque from the respective power rollers 6, 6 (refer to FIGS. 5 to 7), the transmission ratio of the whole continuously variable transmission apparatus is displaced towards the reduce-speed side in order to increase the rotational speed of the engine to decrease the torque TIN with which the input shaft 1 is driven. Due to this, the speed ratio of the toroidal-type continuously variable transmission unit 25 is changed to the increase-speed side.
However, in a state in which the vehicle is stopped with the brake pedal being depressed (the rotational speed of the output shaft=0), the control of the speed ratio of the toroidal-type continuously variable transmission unit 25 is implemented within a range where the change in speed ratio can be absorbed by a slip generated in the interior of the toroidal-type continuously variable transmission unit 25, or a slip generated at the abutment portions (traction portions) between the inner surfaces of the input and output side disks 2, 5a and the circumferential surfaces of the respective power rollers 6, 6 (refer to FIGS. 5 to 7). Consequently, a permissible range where the speed ratio can be regulated is limited to a range where no unbearable force is applied to the abutment portions, and hence, when compared with the low-speed running case, the permissible range becomes limited.
For example, in the event that, with the target value TCVU2 existing at the point a in FIG. 10, the measured value TCVU1 exists at a point b in the same figure, it indicates the existence of a state in which the measured value TCVU1 has the deviation in the direction in which the respective input side disks 2, 2 receive the torque from the respective power rollers 6, 6. Then, the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 is changed to the increase-speed side, so that the speed ratio eCVT of the whole continuously variable transmission apparatus (T/M) is changed to the reduce-speed side. In association with this, the rotational speed of the engine is increased so as to decrease the torque thereof. On the contrary, in the event that the measured value TCVU1 exists at a point c in the same figure, it indicates the existence of a state in which the measured value TCVU1 has a deviation in a direction in which the respective input side disks 2, 2 apply the torque to the respective power rollers 6, 6. In this case, on the contrary to the case described above, the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 is changed to the reduce-speed side, so that the speed ratio eCVT of the whole continuously variable transmission apparatus (T/M) is changed to the increase-speed side. In association with this, the rotational speed of the engine is decreased so as to increase the torque thereof.
Thus, the operations that have been described above are repeatedly performed until the torque TCVU1 passing through the toroidal-type continuously variable transmission unit 25 that is obtained from the differences in pressure coincides with the target value. Namely, in the event that the torque TCVU1 passing through the toroidal-type continuously variable transmission unit 25 cannot be made to coincide with the target value through a single control of the transmission ratio of the toroidal-type continuously transmission unit 25, the operations mentioned above will be carried out repeatedly. As a result, the torque TIN with which the engine rotatably drives the input shaft 1 can be made to approach a value which makes the torque TCVU1 passing through the toroidal-type continuously variable transmission unit 25 coincide with the target value TCVU2. Note that these operations are implemented automatically and within a short period of time by commands from a microcomputer incorporated in a control unit for the continuously variable transmission apparatus.
In addition, FIG. 11 shows relations among a ratio (a left-hand side axis of ordinate) between the torque TCVU passing through the toroidal-type continuously variable transmission unit 25 and the torque TIN with which the engine rotatably drives the input shaft 1 the speed ratio eCVT (a right-hand side axis of ordinate) of the whole continuously variable transmission apparatus and the speed ratio eCVU (an axis of abscissa) of the toroidal-type continuously variable transmission unit 25. A solid line a denotes a relation between the ratio of the passing torque TCVU and the driving torque TIN and the speed ratio eCVT of the whole continuously variable transmission apparatus and a broken line b denotes a relation between both the speed ratios eCVT, eCVU, respectively. According to this construction, with the speed ratio eCVT of the whole continuously variable transmission apparatus being regulated to a predetermined value, the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 is regulated so that the torque TCVU1 which passes through the toroidal-type continuously variable transmission unit 25 is regulated to a target value (TCVU2) which is represented by a point on the solid line a.
In the case of the aforesaid construction, as has been described above, a control for regulation of the torque TCVU1 which actually passes through the toroidal-type continuously variable transmission unit 25 to the point on the solid line a which is the target value TCVU2 is implemented in two stages; that is, by roughly controlling the rotational speed of the engine or controlling the rotational speed of the engine to a rotational speed where the target value TCVU2 is likely to be obtained, and thereafter by controlling the transmission ratio of the toroidal-type continuously variable transmission unit 25 in accordance with the rotational speed so controlled. Due to this, the torque TCVU1 which actually passes through the toroidal-type continuously variable transmission unit 25 can be regulated to the target value TCVU2 without causing an overshoot (and a hunting in association therewith) that is inherent in the conventional methods, or with an overshoot (and a hunting in association therewith) being suppressed to a low level that causes no practical problem, in case the occurrence of overshoot and hunting in association therewith cannot be avoided.
In addition, as has been described before, with the vehicle being stopped by depressing the brake pedal, a driving force (torque) is applied to the output shaft 38a (FIG. 8) based on the slip generated in the interior of the toroidal-type continuously variable transmission unit 25. It is considered that the magnitude of the torque is set to a value which matches a creeping force that is generated in a general automatic transmission apparatus provided with a conventionally propagated torque converter. This is because drivers who are accustomed to operating the general automatic transmission apparatus can avoid having to feel an abnormal feeling. In addition, the direction of the torque is determined by operating positions of an operation lever provided in the vicinity of the driver's seat. When the operation lever selects a forward position (a D range), the torque is applied to the output shaft 38a in a forward direction, whereas when the operation lever selects a reverse position (an R range), the torque is applied to the output shaft 38a in a reverse direction.
Next, a circuit for controlling the speed ratio of the toroidal-type continuously variable transmission unit 25 so that the torque TCVU1 that actually passes through the toroidal-type continuously variable transmission unit 25 coincides with the target value TCVU2 will be described by reference to FIG. 12.
A hydraulic fluid is freely supplied to or discharged from the pair of hydraulic pressure chambers 24a, 24b which constitute the hydraulic actuator 10 for displacing the trunnion 7 in axial directions (in vertical directions in FIG. 12) of the pivot shafts 9, 9 (refer to FIG. 6) through a control valve 12. A sleeve 14 which constitutes the control valve 12 is adapted to be freely displaced in axial directions by a stepping motor 13 via a rod 53 and a link arm 54. In addition, a spool 15 which constitutes the control valve 12 is brought into engagement with the trunnion 7 via a link arm 19, a precess cam 18 and a rod 17 and is adapted to be freely displaced in the axial directions in association with the axial displacement and the swinging displacement of the trunnion 7. This construction is basically the same as that of a control mechanism for controlling the transmission ratio of the toroidal-type continuously variable transmission unit that has been conventionally known.
In particular, with the above construction, the sleeve 14 is, in addition to being driven by the stepping motor 13, driven by a hydraulic differential pressure cylinder 55, as well. Namely, a distal end portion of the rod 53 connected to a proximal end portion of the sleeve 14 is pivotally supported on an intermediate portion of the link arm 54, and pins that are provided at output portions of the stepping motor 13 and the differential pressure cylinder 55 are brought into engagement with elongate holes formed in both end portions of the link arm 54, respectively. In the event that the pin in the elongate hole formed in one end portion of the link arm 54 is pushed and pulled, the pin in the elongate hole in the other end portion of the link arm 54 constitutes a fulcrum. According to this construction, the sleeve 14 is adapted to be displaced axially not only by the stepping motor 13 but also by the differential pressure cylinder 55. In the construction, the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 is designed to be regulated according to the torque TCVU that passes through the toroidal-type continuously variable transmission unit 25 through the displacement of the sleeve 14 by the differential pressure cylinder 55. With the above construction, the speed ratio eCVU of the toroidal-type continuously variable transmission unit 25 is designed to be regulated through the displacement of the sleeve 14 by the differential pressure cylinder 55 according to the torque TCVU that passes through the toroidal-type continuously variable transmission unit 25.
Due to this, with the construction, hydraulic pressures which are different from each other are designed to be freely introduced into a pair of hydraulic pressure chambers 56a, 56b provided in the differential pressure cylinder 55 through a compensating control valve 57. Hydraulic pressures that are introduced into the respective hydraulic pressure chambers 56a, 56b are determined based on a differential pressure ΔP between hydraulic pressures PDOWN, PUP which are applied to the interiors of the pair of hydraulic pressure chambers 24a, 24b which constitute the actuator 10 and a differential pressure ΔPo between output pressures of a pair of electromagnetic valves 58a, 58b which are adapted for regulating the opening of the compensating control valve 57. Namely, the opening and closing of both the electromagnetic valves 58a, 58b is operated by a controller, not shown, such that the differential pressure ΔPo between the output pressures of both the electromagnetic valves 58a, 58b becomes a target differential pressure which corresponds to the target torque TCVU2 of the toroidal-type continuously variable transmission unit 25 and is controlled based on output signals outputted from this control unit. Consequently, a force corresponding to the differential pressure ΔP between the hydraulic pressures applied to the interiors of the hydraulic pressure chambers 24a, 24b of the actuator 10 and a force opposing the force, that is, the differential pressure ΔPo between the output pressures of the electromagnetic valves 58a, 58b which is the target differential pressure which corresponds to the target torque TCVU2 are applied to a spool 59 which constitutes the compensating control valve 57.
In the event that the torque TCVU1 that actually passes through the toroidal-type continuously variable transmission unit 25 coincides with the target torque TCVU2, that is, in the event that a difference ΔT between the passing torque TCVU1 and the target torque TCVU2 is 0, the force corresponding to the differential pressure ΔP between the hydraulic pressures applied to the interiors of the hydraulic pressure chambers 24a, 24b of the actuator 10 and the force corresponding to the differential pressure ΔPo between the output pressures of the electromagnetic valves 58a, 58b balance with each other. Due to this, the spool 59 which constitutes the compensating control valve 57 is positioned at an intermediate position, the pressures applied to the hydraulic pressure chambers 56a, 56b of the differential pressure chamber 55 become equal. In this state, a spool 60 of the differential pressure cylinder 55 is positioned at an intermediate position, and the transmission ratio of the toroidal-type continuously variable transmission unit 25 remains unchanged (no compensation is implemented).
On the other hand, in the event that there is caused a difference between the torque TCVU1 that actually passes through the toroidal-type continuously variable transmission unit 25 coincides and the target torque TCVU2, the balance is broken between the force corresponding to the differential pressure ΔP between the hydraulic pressures applied to the interiors of the hydraulic pressure chambers 24a, 24b of the actuator 10 and the force corresponding to the differential pressure ΔPo between the output pressures of the electromagnetic valves 58a, 58b. Then, the spool 59 which constitutes the compensating control valve 57 is displaced axially according to the magnitude and direction of the difference ΔT between the passing torque TCVU1 and the target torque TCVU2, and an appropriate hydraulic pressure corresponding the magnitude and direction of the difference ΔT is then introduced into the interiors of the hydraulic pressure chambers 56a, 56b of the differential pressure cylinder 55. Then, the spool 60 of the differential pressure cylinder 55 is displaced axially, and in conjunction with this, the sleeve 14 which constitutes the control valve 12 is displaced axially. As a result, the trunnion 7 is displaced in the axial direction of the pivot shafts 9, 9, whereby the speed ratio of the toroidal-type continuously variable transmission unit 25 is changed (a compensation is implemented). Note that the direction and amount in which the transmission ratio is changed as has been described above are as has been described with respect to FIGS. 10 to 11. In addition, the amount in which the speed ratio of the toroidal-type continuously variable transmission unit 25 is changed as has been described above, or the amount in which the speed ratio thereof is compensated (the compensation amount of speed ratio) is sufficiently small relative to the width of the speed ratio of the toroidal-type continuously variable transmission unit 25. Due to this, the stroke of the spool 60 of the differential pressure cylinder 55 is made sufficiently smaller than the stroke of the output portion of the stepping motor 13.
With the continuously variable transmission apparatus that has been constructed as has been described above, when the vehicle is being stopped or is driving at extremely low speeds, it is possible to strictly regulate the speed ratio of the toroidal-type continuously variable transmission unit incorporated in the continuously variable transmission apparatus. However, in order to enable the implementation of a highly accurate control at low cost so as to realize a running condition which makes the driver feel no abnormal feeling, it has to be designed such that the rotational speed of the output shaft 38, 38a is obtained by inexpensive sensors. For example, it is preferred that the torque applied to the drive wheels at the time of so-called creep driving in which the vehicle is driven at extremely low speeds with both the acceleration and brake pedals being released is such that the torque is made to become larger (maximum) with the vehicle being stationary, whereas the torque is made to drastically decrease as the vehicle speed increases. Furthermore, in a state in which, while the vehicle is being stopped on a climbing up slope, the vehicle is forced to be reversed due to an insufficient torque being applied to the drive wheels, it is preferred that on detecting a low-speed rotation of the output shaft 38, 38a in a reverse direction, the torque transmitted from the output shaft 38, 38a to the drive wheels be increased.
While the continuously variable transmission apparatus that is constructed as has been described above is such as to satisfy the aforesaid demand, in order to make that happen, the rotational speed of the output shaft 38, 38a while rotating at extremely low speeds or the stationary state thereof has to be grasped with accuracy. However, a rotation sensor which can grasp the rotational speed of the output shaft 38, 38a while rotating at extremely low speeds or the stationary state thereof with accuracy increases the costs. A so-called active-type rotation sensor is known in which magnetism detection elements such as hall elements and magnetic resistance elements are combined with a permanent magnet, but the rotation sensor of the type is expensive. Further, in a case of that rotational speed is extremely low, it can not be avoided that an response is slow. In contrast to this, since a so-called passive-type rotation sensor is more inexpensive than the active-type rotation sensor in which an electric current which changes in a sine-wave-like fashion is induced in a coil wound around the periphery of a pole piece through which magnetic flux generated from a permanent magnet is conducted, when it is incorporated in the continuously variable transmission apparatus, the attainment of a reduction in cost can be attempted. However, with the passive-type rotation sensor, in case the rotation speed of a portion to be detected is slow, the voltage of electric current induced in the coil becomes low, and the rotation speed of the relevant portion cannot be detected.
Consequently, the passive-type rotation sensor is not appropriate for the rotational speed detection sensor for the output shaft 38, 38a of the continuously variable transmission apparatus. In addition, not only with this passive-type rotation sensor but also with the active-type rotation sensor, neither of them can solely identify the rotational direction of the portion being detected. Due to this, the reverse movement of the automotive vehicle on the climbing up slope that has been described previously cannot be identified. While, with a plurality of rotational sensors being combined together, such a movement can be identified, increases in costs and installation space cannot be avoided.